System and method to produce liquefied natural gas

ABSTRACT

A small to mid-scale liquefied natural gas production system and method is provided. The disclosed liquefied natural gas production system employs at least one heat exchanger, three turbine/expanders and at least three refrigerant compression stages. The expansion ratio of one turbine/expander is appreciably lower than the expansion ratio of the other turbine/expanders such that the temperature of the exhaust stream from the turbine/expander with the lower expansion ratio is above the critical point temperature of the compressed natural gas containing feed stream but colder than about −15° C. The present system and method may be configured using either a single nitrogen-based expansion refrigerant circuit or two separate refrigerant circuits wherein the turbine/expander with the lowest expansion ratio is contained within a separate refrigeration circuit from the other two turbine/expanders with the higher expansion ratios.

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of and priority to U.S. provisionalpatent application Ser. No. 63/270,186 filed Oct. 21, 2021.

TECHNICAL FIELD

The present invention relates to production of liquefied natural gas(LNG), and more particularly, to a small or mid-scale liquefied naturalgas production systems and methods using a nitrogen based refrigerantthat employs at least three turbine/expanders and at a plurality ofrefrigerant compression stages and employing either a single refrigerantcircuit or two separate refrigerant circuits.

BACKGROUND

Demand for liquified natural gas production in applications related toenergy infrastructure, transportation, heating, power generation israpidly increasing. The use of liquified natural gas as a lower cost,alternative fuel also allows for a potential reduction in carbonemissions and other harmful emissions such as nitrogen oxides (NOx),sulphur oxides (SOx), and particulate matter which are generallyrecognized as detrimental to air quality. As a result of this demand, atrend has emerged for construction and operation of lower capacityliquified natural gas production systems built in regions whereattractive sources of low cost natural gas or methane biogas areavailable and/or where there is a current demand for liquified naturalgas, or the demand is expected to grow over time.

Small-scale to mid-scale liquified natural gas opportunities includevarious energy applications such as oil well seeding or boil-off gasre-liquefaction, integrated CO₂ extraction and natural gas liquefaction,utility sector applications such as peak-shaving or emergency reserves,liquified natural gas supply at compressed natural gas filling stations,and transportation applications including marine transportationapplications, off-road transportation applications, and even on-roadfleet transportation uses. Other small-scale or mid-scale liquifiednatural gas opportunities might include liquified natural gas productionfrom biogas sources such as landfills, farms, industrial/municipal wasteand wastewater operations.

Most conventional small-scale or mid-scale liquified natural gasproduction systems target a production of between 100 mtpd and 500 mtpdof liquified natural gas and higher. Many of these liquefaction systemsemploy mechanical refrigeration or a nitrogen-based gas expansionrefrigeration cycles to cool to the natural gas feed to temperaturesrequired for natural gas liquefaction. Use of nitrogen-based gasexpansion refrigeration cycles are the preferred technology for smallscale applications due to simplicity, safety, ease of operation,turndown, dynamic responsiveness and maintenance.

The current market for such small scale natural gas liquefaction systemsusing nitrogen-based gas expansion refrigeration cycles is dominated bythe sale of equipment. Even though many recent opportunities are drivenby environmental considerations, minimizing the installed cost of suchnatural gas liquefaction systems is a dominant factor in theliquefaction process design. When designing natural gas liquefactioncycles and liquefaction systems, capital costs and operationalefficiency must be balanced. Such design decisions are highly dependenton site-specific variables, including natural gas feed quality as wellas the intended applications and transport of the liquified natural gasproduct.

In a conventional high-pressure natural gas liquefaction systememploying nitrogen-based gas expansion refrigeration cycle with dualexpansion, such as that shown in FIG. 1 , there exists a need to improvethe thermal efficiency of such systems. The use of only twoturbine/expanders and the condensing profile of natural gas result inmeaningful divergences in the heat exchanger composite curves. Couplingthe turbine/expanders to one or more compression stages via an integralgear machine or ‘compander’ further complicates efforts to improvethermal performance. Specifically, one cannot simply manipulate turbineexpansion ratio, flow and thermal positioning independent a simultaneousconsideration of the turbomachinery performance and any such additionalturbomachinery-complexity would require a power reduction commensuratelylarge to offset such additional capital.

Another limiting aspect of the conventional natural gas liquefactionsystem and process depicted in FIG. 1 is found with respect to thetemperature levels served by each of the turbine/expanders. Since thecold turbine/expander provides the subcooling duty necessary to preventany meaningful loss of product upon depressurization, the exit state islargely fixed by the cold-end delta temperature (CEDT) of the heatexchanger and the condition of saturation (which minimizes unit powerconsumption). The cold turbine/expander inlet state is defined by anarrow range of temperature in which the coldest portion of thecomposite curves can be made to roughly match. As the inlet temperatureto the cold turbine/expander approaches the pseudo dew-point inflectiontemperature of natural gas, it becomes impossible for the warmingexhaust flow to match the subcooling curve of natural gas. Given theseconsiderations, and the parallel arrangement, the pressure ratios arelargely fixed and/or limited by the cold turbine/expander operation.

The conventional two turbine/expander liquefaction system shown in FIG.1 also exhibits a highly skewed distribution of refrigeration. Since thewarm turbine/expander in such conventional natural gas liquefactionsystems discharges below the critical point temperature of the naturalgas (i.e. −82.6° C.), its flow absorbs much of the duty associated withprecooling the refrigerant and natural gas flows as well as the duty ofNG pseudo-condensation. In the conventional natural gas liquefactionsystem and process depicted in FIG. 1 , the warm turbine/expanderaccounts for about 69% of the recycle refrigerant flow and suppliesabout 83% of the delivered refrigeration. Consequently, the absorbedpower of pinion #2 which couples the cold turbine/expander to adownstream compression stage is substantially higher than that of pinion#1 which couples the warm turbine/expander to an upstream compressionstage. This arrangement complicates both the design of theturbomachinery as well as the ability of the process to fully utilizethe capacity of any given ‘compander’ frame.

What is needed therefore, is a natural gas liquefaction system andprocess that provides a more equitable distribution of power to theindividual pinions and which exhibits an outsized capitalized powerbenefit relative to the conventional two turbine/expander liquefactionsystems with limited added capital expense.

Another natural gas liquefaction system that discloses a threeturbine/expander based natural gas liquefaction cycle is disclosed inU.S. Pat. No. 5,768,912 (Dubar). In that prior art disclosure, threebooster loaded nitrogen expanders are disposed in series and theresulting efficiencies of this Dubar based three turbine/expanderliquefaction arrangement is less than ideal resulting in additionalcapital costs without the corresponding reduction in power and operatingcosts.

Thus, what is also needed are improvements in the overall design andperformance of such natural gas liquefaction systems and processes withthe objective of minimizing the heat exchange liquefactioninefficiencies while facilitating turbomachinery design. In this way,power consumption can be minimized. This goal of minimizing the heatexchange liquefaction inefficiencies is critical to achieving meaningfulperformance improvements.

SUMMARY OF THE INVENTION

The present invention may be characterized as a natural gas liquefactionsystem comprising: a refrigeration circuit and an integral gear machine.The refrigeration circuit includes: a natural gas liquefaction system,comprising: (i) at least one heat exchanger configured to liquefy andsubcool a compressed natural gas containing feed stream via indirectheat exchange with a refrigerant stream; (ii) three or moreturbine/expanders configured to expand portions of the refrigerantstream to produce at least three exhaust streams that are directed tothe at least one heat exchanger to liquefy and subcool the natural gascontaining feed stream via indirect heat exchange and exit the at leastone heat exchanger as one or more warmed recycle streams; and (iii) atleast three refrigerant compression stages including an upstreamrefrigerant compression stage and a pair of downstream refrigerantcompression stages arranged in parallel, wherein the three refrigerantcompression stages are configured to compress the warmed recyclestreams. The integral gear machine includes a drive assembly, a bullgear, and at least three pinions arranged to drive the at least threerefrigerant compression stages and/or for receiving work produced by theat least three turbines/expanders. All three pinions are configured tobe net absorbers of power from the drive assembly of the integral gearmachine and the power is distributed to these three pinions in generallyequal or roughly equal proportions of between 30% and 40% of the totalpower to each of the three pinions.

More specifically, the three or more turbines/expanders furthercomprise: a cold turbine/expander configured to expand a cold portion ofthe refrigerant stream and produce a cold exhaust that is also recycledto the upstream refrigerant compression stage; a first warmturbine/expander configured to expand a first warm portion of therefrigerant stream and produce a first warm exhaust to be recycled tothe downstream refrigerant compression stages; and a second warmturbine/expander configured to expand a second warm portion of therefrigerant stream and produce a second warm exhaust to be recycled tothe downstream refrigerant compression stages.

The present invention may also be characterized as a natural gasliquefaction system comprising at least one heat exchanger configured toliquefy and subcool a compressed natural gas containing feed stream viaindirect heat exchange with a nitrogen-based refrigerant stream from afirst refrigeration circuit and a secondary refrigerant streamtraversing a secondary refrigeration circuit. The first refrigerationcircuit includes at least two turbine/expanders configured to expandportions of the nitrogen-based refrigerant stream to produce one or moreexhaust streams that are directed to the at least one heat exchanger toliquefy and subcool the natural gas containing feed stream via indirectheat exchange and exit the at least one heat exchanger as one or morewarmed recycle streams. The first refrigeration circuit also includes atleast two primary refrigerant compression stages including an upstreamrefrigerant compression stage and a serially arranged downstreamrefrigerant compression stage configured to compress the warmed recyclestreams. The second refrigeration circuit includes at least oneturbine/expander configured to expand portions of the secondaryrefrigerant stream to produce a secondary exhaust stream that isdirected to the heat exchanger and exits the heat exchanger as a warmedsecondary recycle stream. The second refrigeration circuit also includesat least one secondary refrigerant compression stage configured tocompress the warmed secondary recycle stream. If used, the secondaryrefrigerant is a different composition than the nitrogen basedrefrigerant stream and preferably a natural gas or other refrigerant,including hydrocarbon based refrigerants.

When using the separate refrigeration circuits, the coldturbine/expander is configured to expand a cold portion of thenitrogen-based refrigerant stream and produce a cold exhaust that isalso recycled to the upstream refrigerant compression stage. The firstwarm turbine/expander is configured to expand a first warm portion ofthe nitrogen-based refrigerant stream and produce a first warm exhaustto also be recycled to the upstream refrigerant compression stage. Thesecond warm turbine/expander is configured to expand a second warmportion of the secondary refrigerant stream and produce a second warmexhaust to be recycled to the secondary refrigerant compression stage inthe second refrigeration circuit.

In all embodiments, the expansion ratio of the secondary warmturbine/expander is lower than an expansion ratio of the coldturbine/expander and lower than an expansion ratio of the warmturbine/expander. The second warm exhaust is above the critical pointtemperature of the natural gas containing feed stream and preferablyless than about −15° C. Preferably, the first warm turbine/expander hasan expansion ratio of between 4.0 and 5.0 and is configured to producethe majority of the turbine work used to produce the refrigerationwhereas the cold turbine/expander also has an expansion ratio of between4.0 and 5.0 and is configured to produce less than 25% of the turbinework used to produce the refrigeration. The second warm turbine/expanderpreferably has an expansion ratio of between 1.5 and 2.5 and isconfigured to produce between about 20% to 35% of the turbine work.

In all embodiments, an integral gear machine comprising a driveassembly; a bull gear; and at least three pinions is configured to drivethe plurality of refrigerant compression stages and for receiving workproduced by the plurality of turbine/expanders. Also, the purified,compressed natural gas feed stream is preferably at a pressure greaterthan the critical pressure of natural gas, and more preferably at apressure between about 50 bar(a) and 80 bar(a). The refrigerant streamis a nitrogen-based refrigerant that preferably comprises more thanabout 80% nitrogen by volume.

BRIEF DESCRIPTION OF THE DRAWINGS

It is believed that the claimed invention will be better understood whentaken in connection with the accompanying drawings in which:

FIG. 1 shows a generalized schematic of the process flow diagram for aconventional two turbine and two refrigerant compression stage naturalgas liquefaction process known in the prior art;

FIG. 2 shows a schematic of the process flow diagram for an embodimentof the present system and method for liquefied natural gas productionusing three turbine/expanders and three refrigerant compression stagesusing a single refrigerant and having two of the three refrigerantcompression stages arranged in parallel; and

FIG. 3 shows a schematic of the process flow diagram for anotherembodiment of the present system and method for liquefied natural gasproduction using three turbine/expanders and three refrigerantcompression stages having two separate refrigerant circuits.

DETAILED DESCRIPTION

The design of high efficiency liquefaction processes that employ gasexpansion to provide the refrigeration necessary to liquefy and subcoola purified and compressed natural gas containing feed stream is theresult of a simultaneous considerations of heat transfer andturbomachinery within the system and/or process. The minimization ofheat transfer irreversibility is achieved when the divergence of thewarming and cooling composite curves (e.g. energy transferred vstemperature) is minimized. Process definition of flows, pressures andtemperatures largely control the resulting composite curves.Turbomachinery efficiency is maximized when the head and flowcharacteristics of the process are consistent with experience-basedoptimums. These optimal designs are often characterized by establishedratios of geometry, flow and head (Ns, Ds). Such considerationsresulting from dimensional similarity are well known to the art of gasprocessing. See, for example, the publication entitled ‘How to SelectTurbomachinery for your Application’ by Kenneth E. Nichols. Theseoptimal turbomachinery conditions are a function of the type of machineunder consideration.

In the present system and method, the use of a plurality of centrifugalturbomachines, and, in particular, three radial inflow turbines, findparticular application. The present system and method requires or atleast contemplates the natural gas containing feed being a purified,compressed natural gas feed stream at a pressure greater than thecritical pressure of natural gas but it may originate from a source ofmethane containing biogas. As used herein, the term purified natural gascontaining feed stream means a natural gas feed stream substantiallyfree of heavy hydrocarbons, carbon dioxide, water, and other impurities.The subsequent and direct liquefaction of a sub-critical natural gasfeed stream results in a composite curve divergence near the dewpoint ofthe mixture. Furthermore, liquefaction of natural gas at pressures lowerthan about 40 bar(a) generally results in a colder level of warmturbine/expander operation which in turn creates a meaningful penalty interms of unit power consumption. To avoid this penalty, the natural gascontaining feed stream is preferably at a pressure above the criticalpressure of the natural gas feed stream, and more preferably betweenabout 50 bar(a) and 80 bar(a).

Yet another advantageous feature of the present system and method toproduce liquefied natural gas is the use of an integral gear machinecomprising a drive assembly, a bull gear, and a plurality of pinionsarranged or configured to drive two or more refrigerant compressionstages and/or for receiving work produced by the threeturbine/expanders. The shaft of the bull gear may also be connected viagears to the driver assembly. At least two of the plurality of pinionsare net absorbers of power from the drive assembly, which can be anelectric motor, a steam turbine, or even a gas turbine. An importantaspect or advantage of this integral gear machine arrangements disclosedherein relates to the specific pairings of turbomachinery on thedifferent pinions in a manner that optimizes the performance of thepresent liquefaction system and method.

The optimization of the turbomachinery starts with a consideration ofturbine/expander efficiency. Any given process definition (e.g.Pressures, Temperatures, and Flows) that results in a feasible heattransfer (liquefaction) design also provides the necessary input, suchas flow and head characteristics, that are necessary to define thenon-dimensional characteristics (Ns, Ds) required to specify componentturbine/expander rotational speed and diameter. It is well establishedthat radial inflow turbines reach peak efficiency with U/Co (i.e. RotorTip Speed/Isentropic Spouting Velocity) values near 0.70. This ratio isalso defined by the following equation [U/Co]=[NsDs]/154. As such,effective process definition will dictate the speed and diameternecessary for the turbine/expander to operate at peak efficiency. Withrespect to gas compression, process definition dictates compressionstage head and the associated turbine/expander on the same piniondictates rotational speed which in turn results in a specific speed. Theabove calculation forms one part of the overall process optimization.More specifically, the optimization is an iterative process involvingprocess definition, turbomachine pairing based upon the abovecalculation and finally a consideration of the integral gear machinepinion power and overall input power limitations.

Conventional small-scale and medium-scale liquified natural gas plantsthat use a nitrogen-based gas expansion as the primary source ofrefrigeration typically employ centrifugal recycle compression stagesfor the refrigerant that are typically driven by an integral gearmachine contained within a common housing that includes a large diameterbull gear with several meshing pinions upon the ends of which thevarious compression impellers are mounted forming the plurality ofrefrigerant compression stages and expansion impellers of theturbine/expanders. The pinions may have differing diameters to bestmatch the speed requirements of the coupled compression impellers. Eachof the multiple compression impellers and turbine/expanders aretypically contained within their own respective housings andcollectively provide several stages of recycle compression andexpansion, as desired.

Linde Inc., a member of the Linde Group of Companies, has also developeda portfolio of integral gear machines or single machines that combinecompression stages and high efficiency radial inflow expanders having upto four pinions in what is referred to as an integral gear ‘bridge’machine or BRIM. Linde's ‘bridge’ machines are conventionally used inhydrogen/syngas plants as well as air separation plants and typicallycome in different frame sizes, for example between about 90 mm and 180mm frame sizes. Design studies have examined applications of the Linde‘bridge’ machines to operatively couple a plurality of radially inflowturbines and centrifugal refrigeration compression stages in a naturalgas liquefaction system. The Linde ‘bridge’ machines come fully packagedor integrated with appropriate PLC controllers, control valves, safetyvalves, oil system, etc. and can be easily outfitted with intercoolersand/or aftercoolers. The hardware constraints and limitations of theLinde ‘bridge’ machines are typically a function of bull gear and driverassembly size. In general, the Linde ‘bridge’ machine drivers pertinentfor the present system and method spans the range of about 4 MW to 20 MWwith associated maximum pinion speeds in the range of 20,000 to 50,000rpm. Furthermore, the maximum power imparted to any given pinion or anygiven turbine-compression stage pairing is preferably limited to lessthan 50% and in some cases to about 35% of the total ‘bridge’ machinedriver power.

Three (3) Turbine LNG Production System with Single RefrigerationCircuit

Turning to FIG. 2 , a schematic of the high-level process flow diagramfor one embodiment of the present system and method for liquefiednatural gas production using three turbine/expanders having a singlenitrogen-based expansion refrigerant circuit is shown. The illustratedrefrigerant circuit includes at least one heat exchanger, twoaftercoolers; three turbine expanders, and three refrigerant compressionstages wherein two of the refrigerant compression stages are arranged inparallel. The illustrated system also includes a three-pinion integralgear machine, a fuel gas circuit, and a post liquefaction conditioningcircuit, having one or more expansion valves and a phase separatorconfigured for separating nitrogen and other light gases from theliquefied and subcooled natural gas stream.

The purified and compressed natural gas containing feed substantiallyfree of heavy hydrocarbons and other impurities and at a feed pressurethat is greater than the critical pressure of natural gas (i.e. above 46bar(a)), preferably at a pressure of between about 50 bar(a) and 80bar(a) and more preferably at a pressure between about 60 bar(a) and 75bar(a) is provided as a feed stream to the depicted natural gasliquefaction system.

As indicated above, the purified, compressed natural gas feed stream isliquefied and subcooled within the heat exchanger(s) via indirect heatexchange against one or more nitrogen-based refrigerant streams to forma subcooled and liquified natural gas stream. The subcooled andliquified natural gas stream is thereafter treated in the postliquefaction conditioning circuit where the subcooled and liquefiednatural gas is reduced in pressure via one or more valves, or a liquidturbine (not shown), and phase separated using a phase separator toseparate nitrogen vapor and other light gases. The resulting liquidnatural gas stream constitutes the liquefied natural gas product.

The primary refrigeration source used in the illustrated natural gasliquefaction system is preferably a nitrogen-based gas expansionrefrigeration circuit, that preferably includes refrigerant stream(s)that comprises more than about 80% nitrogen by volume. In suchillustrated refrigeration circuit, the refrigerant is compressed in aplurality of refrigerant compression stages, namely an upstreamrefrigerant compression stage and two downstream refrigerant compressionstages arranged in parallel with appropriate intercooling and/oraftercooling used to offset the temperature increases caused by the heatof compression. Such aftercooling may be accomplished by way of indirectcontact with air, cooling water, chilled water or other refrigeratingmedium or combinations thereof. The compressed refrigerant stream isthen further cooled in the at least one heat exchanger(s) and directedto one or more turbine/expanders configured to expand the compressedrefrigerant streams to generate refrigeration.

While the embodiment of FIG. 2 depicts a single heat exchanger havingmultiple warming passages and multiple cooling passages. Alternatively,the at least one heat exchanger can include multiple heat exchangers ormultiple heat exchange cores with a first heat exchanger, or first heatexchange core configured for liquefying the natural gas feed stream anda second heat exchanger or second heat exchange core configured forcooling other streams, such as pre-cooling a portion of the refrigerantstream or perhaps even pre-cooling the natural gas feed stream. Any suchsecond heat exchanger or second heat exchange core would preferablyachieve such pre-cooling with the exhaust stream from the second warmturbine, discussed below.

Specifically, a first portion of the compressed refrigerant stream issubstantially cooled in the heat exchanger and directed to a coldturbine/expander as a cold portion of the refrigerant stream. A secondportion of the compressed refrigerant stream is partially cooled andexits the heat exchanger at an intermediate warmer temperature as afirst warm portion which is then directed to a first warmturbine/expander. A third portion of the compressed refrigerant streamis also partially cooled and exits the heat exchanger as a second warmportion of the compressed refrigerant stream having a temperature warmerthan the intermediate warmer temperature. The second warm portion of thecompressed refrigerant stream is then directed to a second warmturbine/expander.

The cold turbine/expander is configured to expand the cold portion ofthe compressed refrigerant stream to produce a cold exhaust stream thatis recycled back to the refrigerant compression stages via one or moreof the plurality of warming passages in the heat exchanger(s). Thepartially cooled first warm portion of the compressed refrigerant streamis expanded in the first warm turbine/expander to produce a first warmexhaust stream that is also recycled to the one or more refrigerantcompression stages via one or more of the plurality of warming passagesin the heat exchanger(s). The partially cooled second warm portion ofthe compressed refrigerant stream is expanded in the second warmturbine/expander to produce a second warm exhaust stream that is alsorecycled to the one or more refrigerant compression stages via one ormore of the plurality of warming passages in the heat exchanger(s).

The inlet pressures of the three turbine/expanders are approximatelyequal but the outlet pressures are different. Specifically, theexpansion ratio of the cold turbine/expander and the first warm turbineexpander are preferably between about 4.0 and 5.0. Using similarexpansion ratios, the cold exhaust and the first warm exhaust may bewarmed in the heat exchanger using the same warming pressure.Alternatively, the cold exhaust and the first warm exhaust may be warmedin independent passages of the heat exchanger(s) and/or may be atdifferent outlet pressures. An important and advantageous feature of thepresent system and method is that the second warm turbine/expander hasan expansion ratio much less than the expansion ratio of the coldturbine/expander and first warm turbine/expander. Preferably, the secondwarm turbine/expander has an expansion ratio of between 1.5 and 2.5 andsince the second warm exhaust is at a pressure greater than the coldexhaust and the first warm exhaust, it should be warmed in anindependent passage of the heat exchanger(s).

Upon exiting the heat exchanger, the warmed cold turbine exhaust and thewarmed first warm turbine exhaust are recycled as a lower pressurerecycle stream to the upstream refrigerant compression stage where thelower pressure recycle stream is compressed with the resultingcompressed recycle stream being aftercooled in the upstream aftercooler.The warmed second warm turbine exhaust is also recycled as a higherpressure recycle stream and is mixed with the aftercooled, compressedrefrigerant stream exiting the upstream refrigerant compression stage.This mixed stream is then split into a first refrigerant stream and asecond refrigerant stream. The first refrigerant stream is directed toone of the parallel downstream refrigerant compression stages, namely afirst downstream refrigerant compression stage while the secondrefrigerant stream is directed to the other of the parallel downstreamrefrigerant compression stages, namely a second downstream refrigerantcompression stage. The preferred split of the first refrigerant streamand second refrigerant stream is roughly between 35% and 45% of the flowis taken as the first refrigerant stream while 55% to 65% of the flow istaken as the second refrigerant stream. The first and second refrigerantstreams are then further compressed in the respective downstreamrefrigerant compression stages, recombined into a further compressedrecycle stream and subsequently cooled in the downstream aftercooler.

In the depicted embodiment, the cold exhaust is at a temperature colderthan −145° C. while the first warm exhaust is at a temperature colderthan −90° C. but warmer than the cold exhaust. The second warm exhaustis at a temperature above the critical point temperature of thecompressed natural gas feed stream and warmer than the first warmexhaust and preferably colder than about −15° C. Also, the distributionof the compressed refrigerant stream between the cold portion, the firstwarm portion, and the second warm portion is such that the first warmturbine/expander is configured to produce over 45% of the turbine workused to produce the refrigeration for the natural gas liquefactionsystem. The cold turbine/expander is configured to produce less than 25%of the turbine work used to produce the refrigeration for the naturalgas liquefaction system while the second warm turbine/expander isconfigured to produce between about 25% to 35% of the turbine work usedto produce the refrigeration for the liquefaction system.

The first warm turbine/expander, the second warm turbine/expander, andthe cold turbine/expander as well as the upstream refrigerantcompression stage and the downstream refrigerant compression stages areoperatively coupled to the integral gear machine. In particular, thefirst downstream refrigerant compression stage and the coldturbine/expander are operatively coupled to the same pinion of theintegral gear machine, identified as the second pinion of the threepinion integral gear machine. Likewise, the second downstreamrefrigerant compression stage and the second warm turbine/expander areoperatively coupled to the same pinion of the integral gear machine,shown as the first pinion. The first warm turbine/expander and theupstream refrigerant compression stage are coupled to yet a differentpinion, shown as the third pinion of the integral gear machine.

A portion of the liquified and subcooled natural gas feed stream may bediverted to the fuel gas circuit. The fuel gas circuit includes one ormore valves configured to expand the diverted portion of the liquifiedand subcooled natural gas stream to a pressure less than about 6.0bar(a). The lower pressure fuel gas stream is then directed to the heatexchanger to subcool the purified, compressed natural gas stream withthe warmed, low pressure fuel gas stream exits the heat exchanger nearambient temperature to be used or stored as fuel gas.

Three (3) Turbine LNG Production System with Separate RefrigerationCircuits

Turning to FIG. 3 , there is shown a schematic of the high-level processflow diagram for another embodiment of the present system and method forliquefied natural gas production using three turbine/expanders and threerefrigerant compression stages. Many of the features, components andstreams associated with the natural gas liquefaction system shown inFIG. 3 are similar or identical to those described above with referenceto the embodiment of FIG. 2 and for sake of brevity will not be repeatedhere. The key differences between the natural gas liquefaction systemshown in FIG. 3 compared to the natural gas liquefaction systemdescribed above with reference to FIG. 2 , is the addition of a separateand distinct refrigeration circuit that includes the second warmturbine/expander, one of the downstream refrigerant compression stages,a second downstream aftercooler, and dedicated cooling and warmingpassages in the at least one heat exchanger. Note that by separating thetwo downstream refrigerant compression stages into separaterefrigeration circuits, the downstream refrigerant compression stagesare no longer arranged in parallel in the embodiment of FIG. 3 .

The natural gas liquefaction system illustrated in FIG. 3 includes afirst refrigerant circuit with the cold turbine/expander, the first warmturbine expander, the upstream refrigerant compression stage, a firstdownstream refrigerant compression stage, the at least one heatexchanger and two aftercoolers. The natural gas liquefaction systemillustrated in FIG. 3 also includes a mixed service integral gearmachine, a fuel gas circuit, and a post liquefaction conditioningcircuit, similar to those described above with reference to FIG. 2 . Thenatural gas liquefaction system of FIG. 3 also includes a separate anddistinct second refrigeration circuit that includes the second warmturbine/expander, one of the downstream refrigerant compression stages,a second downstream aftercooler, and dedicated cooling and warmingpassages in the at least one heat exchanger.

As indicated above, the purified and compressed natural gas feed is at afeed pressure that is greater than the critical pressure of natural gasand preferably at a pressure of between about 50 bar(a) and 80 bar(a).The primary refrigerant source in the first refrigeration circuit ispreferably a stream that comprises more than about 80% nitrogen byvolume while the secondary refrigerant source in the secondrefrigeration circuit has a different composition than thenitrogen-based refrigerant and is preferably a stream of natural gas oranother hydrocarbon based refrigerant.

As detailed in FIG. 3 , the warmed cold exhaust and the warmed firstwarm exhaust exiting the heat exchanger are recycled as a lower pressurenitrogen recycle stream to the upstream refrigerant compression stage.The compressed refrigerant stream exiting the upstream refrigerantcompression stage is then cooled in the upstream aftercooler. Theaftercooled stream is then directed to the first downstream refrigerantcompression stage where it is further compressed as further compressedrecycle stream and aftercooled in downstream aftercooler. The cooled,further compressed nitrogen recycle stream is then directed to the atleast one heat exchanger(s) where it is cooled to the appropriate inlettemperatures for the first warm turbine/expander and the coldturbine/expander.

The warmed second warm exhaust is also recycled as a higher pressurerecycle stream and directed to the second downstream refrigerantcompression stage where it is further compressed to form the secondaryrefrigerant stream which is then cooled in the third aftercooler. Thesecondary refrigerant stream is then partially cooled in the heatexchanger(s) and directed to the second warm turbine configured toexpand the secondary refrigerant stream to generate refrigeration.

In this embodiment, the cold exhaust is also at a temperature colderthan about −145° C. while the first warm exhaust is at a temperaturecolder than −90° C. but warmer than the cold exhaust. The second warmexhaust is at a temperature above the critical point temperature of thecompressed natural gas feed stream and warmer than the first warmexhaust and preferably colder than about −15° C. In this embodiment thefirst warm turbine/expander is configured to produce over 60% of theturbine work used to produce the refrigeration for the natural gasliquefaction system while the cold turbine/expander is configured toproduce less than 20% of the turbine work used to produce therefrigeration for the natural gas liquefaction system. The second warmturbine/expander in the separate refrigeration circuit is preferablydesigned or configured to produce between about 20% to 30% of theturbine work used to produce the refrigeration for the liquefactionsystem.

The first warm turbine/expander, the second warm turbine/expander, andthe cold turbine/expander as well as the upstream refrigerantcompression stage and the downstream refrigerant compression stages areoperatively coupled to the integral gear machine. In particular, thefirst downstream refrigerant compression stage and the coldturbine/expander are operatively coupled to the same pinion of theintegral gear machine, identified as the second pinion of the threepinion integral gear machine. Likewise, the first warm turbine/expanderand the upstream refrigerant compression stage are coupled to yet adifferent pinion, shown as the third pinion of the integral gearmachine. Lastly, the secondary refrigerant compression stage and thesecond warm turbine/expander of the second refrigeration circuit areoperatively coupled to the same pinion of the integral gear machine,shown as the first pinion.

Examples of LNG Production

A number of computer simulations were run to characterize theperformance of the present natural gas liquefaction system andprocesses. In one such computer simulation, referred to as Case 1, anatural gas liquefaction system designed to produce 175 metric tonnesper day of liquefied natural gas at 164.4° C. and 1.5 bar(a) from acompressed, purified natural gas feed stream at a pressure of about 68bar(a) and a temperature of about 30° C. was evaluated using thearrangement disclosed above with reference to FIG. 2 .

Table 1A provides the work distribution in this example using theembodiment of the three pinion integral gear machine used in the threeturbine/expander and two refrigerant compression stage systemschematically depicted in FIG. 2 . Similarly, Table 1B provides theprocess flow and refrigerant stream characteristics for this exampleusing the same FIG. 2 embodiment of the three turbine/expander and tworefrigerant compression stage natural gas liquefaction system.

TABLE 1A Net FIG. 2 Power Power Power Pinion# Service #1 (kw) Service #2(kW) (kw) Pinion #1 N2 Comp #CB1 17886 N2 - Cold Turbine −200 860 Pinion#2 N2 Comp #CB2 1119 N2 - 1^(st) Warm Turbine −928 919 Pinion #3 N2 Comp#CB3 1545 N2 - 2^(nd) Warm Turbine −496 1049

TABLE 1B Refrigerant FIG. 2 Temp Flow Stream Description Stream# (° C.)(% of Total) CB1 Inlet M06 33.50 64.0% CB2 Inlet M12 34.75 42.0% CB3Inlet M12A 34.75 58.0% Cold Turbine Inlet M03 −114.75 18.9% Cold TurbineExhaust M04 −166.90 18.9% 1^(st) WT Inlet R02 −40.35 45.1% 1^(st) WTExhaust R03 −114.65 45.1% 2^(nd) WT Inlet S02 9.47 36.0% 2^(nd) WTExhaust S03 −38.00 36.0% Lower Pressure Recycle M06 33.50 64.0% HighPressure Recycle S04 33.50 36.0% Upstream Aftercooler M10 36.00 64.0%Downstream Aftercooler M15 36.00 100.0%

In the Case 1 simulation, the speed of the cold turbine/expander is thevariable that constrains the process cycle and, in this example,approaches a speed of about 40,000 rpm. Note that all other threepinions are net absorbers of power from the drive assembly of theintegral gear machine and the power is distributed to these threepinions in generally equal or roughly equal proportions or 37.1%, 32.5%and 30.4%. Note, however, that the upstream refrigeration compressionstage is designed to compress about 64% of the refrigerant and thiscompressed refrigerant is mixed or combined with the higher pressurerecycle stream which contains the remaining 36% of the refrigerant. Thedownstream refrigerant compression stages arranged in parallel are thusdesigned to further compresses the entire refrigerant stream.

The distribution of the fully compressed refrigerant stream between thecold turbine/expander, the first warm turbine/expander, and second warmturbine/expander in this Case 1 example is such that the first warmturbine/expander is configured to receive almost 49.1% of the compressedrefrigeration stream and expands the stream from an inlet pressure of48.1 bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratioof 4.12. The cold turbine/expander, on the other hand receives only18.9% of the compressed refrigeration stream and expands the stream froman inlet pressure of 47.75 bar(a) to an outlet pressure of 11.78 bar(a)or an expansion ratio of 4.05 while second warm turbine/expanderreceives about 36% of the compressed refrigeration stream and expandsthe stream from an inlet pressure of 48.3 bar(a) to an outlet pressureof 24.15 bar(a) or an expansion ratio of 2.0.

As indicated above, designs of small to mid-scale natural gasliquefaction cycles and liquefaction systems, there are numeroustrade-offs between capital costs and operational efficiencies that mustbe made. The natural gas liquefaction system shown in FIG. 2 andoperated in a manner similar to the example of Case 1 is among a verygood compromise of thermal performance, capital cost and accommodationof turbomachinery constraints.

In another computer simulation, referred to as Case 2, a natural gasliquefaction system designed to produce 175 metric tonnes per day ofliquefied natural gas from a compressed, purified natural gas feedstream at a pressure of about 68 bar(a) and a temperature of about 30°C. was evaluated using the three turbine/expander and three refrigerantcompression stage arrangement disclosed in FIG. 3 . Table 2A providesthe work distribution for the example of Case 2 using the embodiment ofthe three pinion integral gear machine used in the threeturbine/expander and three refrigerant compression stage systemschematically depicted in FIG. 3 while Table 2B provides the processflow and refrigerant stream characteristics for the threeturbine/expander and three refrigerant compression stage natural gasliquefaction system of FIG. 3 .

TABLE 2A Net FIG. 3 Power Power Power Pinion# Service #1 (kw) Service #2(kW) (kw) Pinion #1 N2 Comp #CB1 1788 N2 - 1^(st) Warm Turbine −928 860Pinion #2 N2 Com3 #CB2 1713 N2 - Cold Turbine −200 1513 Pinion #3 NGComp #CB3 761 NG - 2^(nd) Warm Turbine −365 396

TABLE 2B Refrigerant FIG. 3 Temp Flow Stream Description Stream# (° C.)(% of Total) CB1 Inlet M07 33.50 N2 - 100.0% CB2 Inlet M12 36.00 N2 -100.0% CB3 Inlet M07A 32.51 NG - 100.0%  Cold Turbine Inlet M03 −114.75N2 - 29.5%  Cold Turbine Exhaust M04 −166.90 N2 - 29.5%  1^(st) WT InletR02 −40.35 N2 - 70.5%  1^(st) WT Exhaust Inlet R03 −114.65 N2 - 70.5% 2^(nd) WT Inlet S02 −4.84 NG - 100.0%  2^(nd) WT Exhaust S03 −38.00 NG -100.0%  Higher Pressure Recycle S04 32.51 NG - 100.0%  Lower PressureRecycle M08 33.50 N2 - 100.0% Upstream Aftercooler M10 36.00 N2 - 100.0%Downstream N2 Aftercooler M15 36.00 N2 - 100.0% Downstream NGAftercooler S01 36.00 N2 - 100.0%

In the Case 2 simulation, the nitrogen-based or primary refrigerationcircuit has the cold turbine/expander on the second pinion as it ispaired with the first downstream refrigeration compression stage whilethe first warm turbine/expander on the first pinion is paired with theupstream refrigeration compression stage. In the natural gas based orsecondary refrigeration circuit, the second warm turbine/expander on thethird pinion is paired with the natural gas refrigeration compressionstage.

Note all three pinions are net absorbers of power from the driveassembly of the integral gear machine but unlike the Case 1 example, thepower distribution to the three pinions is not uniform. Specifically,the second pinion coupling the cold turbine/expander and the firstdownstream refrigerant compression stage absorbs 55% of the power whilethe first pinion coupling the first warm turbine/expander and theupstream refrigerant compression stage absorbs 31% of the power and thethird pinion coupling the natural gas second warm turbine/expander andthe natural gas compression stage absorbs just over 14% of the power.

The first warm turbine/expander is configured to expand the nitrogenbased or primary refrigerant stream from an inlet pressure of 48.10bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratio ofabout 4.12. The cold turbine/expander also expands the primaryrefrigerant stream from an inlet pressure of 47.75 bar(a) to an outletpressure of 11.78 bar(a) or an expansion ratio of 4.05. In the secondaryrefrigeration circuit, the second warm turbine/expander expands thenatural gas or secondary refrigerant stream from an inlet pressure of67.10 bar(a) to an outlet pressure of 39.5 bar(a) or an expansion ratioof about 1.7.

While the present natural gas liquefaction systems and methods have beendescribed with reference to several preferred embodiments, it isunderstood that numerous additions, changes, and omissions can be madewithout departing from the spirit and scope of the present inventions asset forth in the appended claims.

What is claimed is:
 1. A natural gas liquefaction system, comprising: arefrigeration circuit comprising: (i) at least one heat exchangerconfigured to liquefy and subcool a compressed natural gas containingfeed stream via indirect heat exchange with a refrigerant stream; (ii)three or more turbine/expanders configured to expand portions of therefrigerant stream to produce at least three exhaust streams that aredirected to the at least one heat exchanger to liquefy and subcool thenatural gas containing feed stream via indirect heat exchange and exitthe at least one heat exchanger as one or more warmed recycle streams;and (iii) at least three refrigerant compression stages including anupstream refrigerant compression stage and a pair of downstreamrefrigerant compression stages arranged in parallel, wherein the threerefrigerant compression stages are configured to compress the warmedrecycle streams; an integral gear machine comprising a drive assembly, abull gear, and at least three pinions arranged to drive the at leastthree refrigerant compression stages and/or for receiving work producedby the at least three turbines/expanders; wherein the three or moreturbines/expanders further comprise: (i) a cold turbine/expanderconfigured to expand a cold portion of the refrigerant stream andproduce a cold exhaust that is also recycled to the upstream refrigerantcompression stage; (ii) a first warm turbine/expander configured toexpand a first warm portion of the refrigerant stream and produce afirst warm exhaust to be recycled to the downstream refrigerantcompression stages; and (iii) a second warm turbine/expander configuredto expand a second warm portion of the refrigerant stream and produce asecond warm exhaust to be recycled to the downstream refrigerantcompression stages; wherein an expansion ratio of the secondary warmturbine/expander is lower than an expansion ratio of the coldturbine/expander and lower than an expansion ratio of the warmturbine/expander.
 2. The natural gas liquefaction system of claim 1,wherein the second warm exhaust is above the critical point temperatureof the compressed natural gas containing feed stream and less than about−15° C.
 3. The natural gas liquefaction system of claim 1, wherein thefirst warm turbine/expander is configured with an expansion ratio ofbetween 4.0 and 5.0 and is further configured to produce over 50% of theturbine work used to produce refrigeration for the natural gasliquefaction system.
 4. The natural gas liquefaction system of claim 1,wherein the cold turbine/expander is configured with an expansion ratioof between 4.0 and 5.0 and is further configured to produce less than20% of the turbine work used to produce refrigeration for the naturalgas liquefaction system.
 5. The natural gas liquefaction system of claim1, wherein the second warm turbine/expander is configured with anexpansion ratio of between 1.5 and 2.5 and is further configured toproduce between about 20% to 35% of the turbine work used to producerefrigeration for the natural gas liquefaction system.
 6. The naturalgas liquefaction system of claim 1, wherein the first warmturbine/expander and the upstream compression stage are operativelycoupled to a first pinion of the at least three pinions, and the coldturbine/expander and one of the pair of downstream compression stagesare operatively coupled to a second pinion of the at least threepinions, and the second warm turbine/expander and another of the pair ofdownstream compression stages are operatively coupled to a third pinionof the at least three pinions.
 7. The natural gas liquefaction system ofclaim 1, wherein all three pinions are net absorbers of power and thepower is distributed to these three pinions in generally equal orroughly equal proportions.
 8. The natural gas liquefaction system ofclaim 1, wherein the compressed natural gas containing feed stream is amethane containing biogas feed stream.
 9. The natural gas liquefactionsystem of claim 1, wherein the compressed natural gas containing feedstream is at a pressure greater than the critical pressure of naturalgas.
 10. The natural gas liquefaction system of claim 1, wherein thecompressed natural gas containing feed stream is at a pressure betweenabout 50 bar(a) and 80 bar(a).
 11. The natural gas liquefaction systemof claim 1, wherein the one or more refrigerant streams comprise morethan about 80% nitrogen by volume.
 12. The natural gas liquefactionsystem of claim 1, wherein the driver assembly is an electric motor, asteam turbine, or a gas turbine.
 13. The natural gas liquefaction systemof claim 1, further comprising a phase separator configured forseparating nitrogen and other light gases from the liquefied andsubcooled natural gas stream.
 14. A natural gas liquefaction system,comprising: at least one heat exchanger configured to liquefy andsubcool a compressed natural gas containing feed stream via indirectheat exchange with a nitrogen-based refrigerant stream and a secondaryrefrigerant stream; a first refrigeration circuit comprising at leasttwo turbine/expanders configured to expand portions of thenitrogen-based refrigerant stream to produce one or more exhaust streamsthat are directed to the at least one heat exchanger to liquefy andsubcool the natural gas containing feed stream via indirect heatexchange and exit the at least one heat exchanger as one or more warmedrecycle streams; and at least two primary refrigerant compression stagesincluding an upstream refrigerant compression stage and a seriallyarranged downstream refrigerant compression stage, wherein therefrigerant compression stages are configured to compress the warmedrecycle streams; a second refrigeration circuit comprising at least oneturbine/expander configured to expand portions of the secondaryrefrigerant stream to produce one or more secondary exhaust streams thatare directed to the at least one heat exchanger to liquefy and subcoolthe natural gas containing feed stream via indirect heat exchange andexit the at least one heat exchanger as a warmed secondary recyclestream; and at least one secondary refrigerant compression stageconfigured to compress the warmed secondary recycle stream; and anintegral gear machine comprising a drive assembly; a bull gear; and atleast three pinions arranged to drive the at least two primaryrefrigerant compression stages, the at least one secondary refrigerantcompression stage, and for receiving work produced by theturbines/expanders in the first refrigeration circuit and the secondrefrigeration circuit; wherein the two or more turbines/expanders in thefirst refrigeration circuit further comprise: a cold turbine/expanderconfigured to expand a cold portion of the nitrogen-based refrigerantstream and produce a cold exhaust that is also recycled to the upstreamrefrigerant compression stage; and a first warm turbine/expanderconfigured to expand a first warm portion of the nitrogen-basedrefrigerant stream and produce a first warm exhaust to be recycled tothe upstream refrigerant compression stage; wherein the at least oneturbine/expander in the second refrigeration circuit further comprises:a second warm turbine/expander configured to expand a second warmportion of the secondary refrigerant stream and produce a second warmexhaust to be recycled to the secondary refrigerant compression stage;and wherein an expansion ratio of the secondary warm turbine/expander islower than an expansion ratio of the cold turbine/expander and lowerthan an expansion ratio of the warm turbine/expander.
 15. The naturalgas liquefaction system of claim 14, wherein the second warm exhaust isabove the critical point temperature of the compressed natural gascontaining feed stream and less than about −15° C.
 16. The natural gasliquefaction system of claim 14, wherein the first warm turbine/expanderand the cold turbine/expander are each configured with an expansionratio of between 4.0 and 5.0 and wherein the second warmturbine/expander is configured with an expansion ratio of between 1.5and 2.5.
 17. The natural gas liquefaction system of claim 14, whereinthe first warm turbine/expander and the upstream compression stage areoperatively coupled to a first pinion of the at least three pinions, andthe cold turbine/expander and the downstream compression stage in thefirst refrigeration circuit are operatively coupled to a second pinionof the at least three pinions, and the second warm turbine/expander andthe secondary refrigerant compression stage are operatively coupled to athird pinion of the at least three pinions.
 18. The natural gasliquefaction system of claim 14, wherein the compressed natural gascontaining feed stream is at a pressure greater than the criticalpressure of natural gas and between about 50 bar(a) and 80 bar(a). 19.The natural gas liquefaction system of claim 14, wherein thenitrogen-based refrigerant comprise more than about 80% nitrogen byvolume and the secondary refrigerant has a different composition thanthe nitrogen-based refrigerant.